The conventional spindle motors used for driving the memory device such as the HDD or a bar code reader, includes a hydrodynamic bearing assembly implementing high rotation in a stable manner for a long effective lifetime. Among various features of the hydrodynamic bearing assembly, since rotational member and stationary member of the hydrodynamic bearing assembly operate without contact to each other, the rotation thereof causes mechanical friction therebetween less than that of another contacting type of the bearing assemblies such as a ball bearing. In comparison with the hydrodynamic bearing assembly using oil for generating dynamic pressure, in particular, the hydrodynamic bearing assembly using gas as fluid has further advantages to reduce the coom caused by shattering lubricant such as oil and grease in addition to the reduced friction.
FIG. 42 shows an exemplary spindle motor with a conventional hydrodynamic bearing assembly. In the drawing, the hydrodynamic bearing assembly comprises, on a base plate 200, a column shaft 201, a sleeve 202 rotatably arranged around the shaft 201 leaving a predetermined gap along the axis direction of the shaft 201 for relative rotation therebetween. The hydrodynamic bearing assembly also comprises a thrust plate 202, which is arranged perpendicular to the shaft 201 and opposes to the bottom surface of the sleeve 202. A radial bearing is formed between an outer surface of the shaft 201 and the inner surface of the sleeve 202. Also, a thrust bearing is formed between the bottom surface of the sleeve 202 and the thrust plate 203. The thrust plate 203 includes grooves 205 for generating thrust dynamic pressure, formed on the surface opposing to the bottom surface of the sleeve 202, as illustrated by a dashed line.
In this specification, the bottom surface opposing to the thrust plate 203 and defining the thrust bearing in cooperation therewith is referred to as a thrust opposing surface. In FIG. 42, one of the end surfaces in the axis direction is the thrust opposing surface 204. A rotor 207 attached with the sleeve 202 can be rotated about the shaft 201 with the sleeve 202. The rotor 207 has a rotor magnet 208 arranged on the inner surface of a skirt 207a of the rotor 207. The rotor magnet 208 opposes to the electromagnet 209 arranged on the base plate 200. In case of the HDD, a plurality of memory media are mounted on the outer surface, also in case of the bar code reader, a polygonal mirror is mounted on the rotor 207, both of which rotate with the rotor 207.
According to the spindle motor constructed as described above, an alternating current supplied to the electromagnet 209 causes the attraction and/or repulsion forces between the electromagnet 209 and the rotor magnet 208. This provides the rotor 207 supporting the rotor magnet 208 with a rotation drive force so that the rotor 207 and the sleeve 202 attached therewith together rotate around the axis of the shaft 201. The rotation causes the relative movement between the shaft 201 and sleeve 202, generating the radial dynamic pressure due to the fluid in the radial bearing. In general, although air is often used for the fluid intervening between the shaft 201 and sleeve 202 when the spindle motor is used in the atmosphere, particular gas or oil may be used as the fluid. In this specification, the intervening object for generating the dynamic pressure is referred to as the “fluid”. The aforementioned rotation also causes the relative movement between the thrust opposing surface 204 of the sleeve 202 and the thrust plate 203, thereby generating another dynamic pressure in a thrust direction due to the grooves 205. To this end, this thrust dynamic pressure allows the rotational member such as sleeve 202 and rotor 207 to rotate about the shaft 201 keeping the rotational member away from the stationary member such as shaft 201 and the base 200.
FIG. 43 shows the thrust grooves 205 formed on the surface of the thrust plate 203 for generating the thrust dynamic pressure in the thrust bearing. As shown, the grooves 205 include a plurality of a spiral groove, each of which is angled at a predetermined angle with the circle on the thrust plate 203, and has a depth in a range of 1 micron through 10 microns. The thrust opposing surface 204 of the sleeve 202 rotates in a direction indicated by the arrow 206 against the grooves 205 so that the fluid such as air is convolved in the grooves 205. The fluid is pressed along the spiral grooves 205 towards the axis due to the viscosity of the fluid during the above-mentioned rotation, hereby to generate the pressure (dynamic pressure). This dynamic pressure operates the thrust opposing surface 204 to push up the rotational member such as sleeve 202. Such bearing assembly, which conducts the fluid from the circumference towards the axis of the thrust bearing assembly to generate the dynamic pressure, is referred to as a “pump-in” bearing assembly. The pump-in bearing assembly is commonly used for the hydrodynamic bearing assembly.
A need has been existed in the market to a compact and lightweight hydrodynamic bearing assembly implementing the rotation at high rate and heavy load in a stable manner. There are some problems to be solved for the hydrodynamic bearing assembly to satisfy such market's needs. Firstly, the rotation should be stable in particular at the high rotation rate. Secondary, the bearing assembly should have a certain rigidity sufficient to bear against the oscillation forces provided from external circumstances. Thirdly, the bearing assembly has to be improved in the activation feature to activate rotation of the rotational member in contact with the stationary member. Fourthly, the bearing assembly should be more compact and lightweight. Details for those problems to be solved will be described hereinafter.
(First Problem)
In order to address the first problem, i.e., to realize the high rotation rate in a stable manner, it is necessary to eliminate a phenomenon, so-called half-whirl. The half-whirl is the phenomenon appeared due to the rotation of sleeve 202 relative to the shaft 201 with a predetermined gap for keeping thereof away to each other. The fluid intervening between the outer surface of the shaft 201 and the inner surface of the sleeve 202 for generating the dynamic pressure causes a continuous pressure distribution therebetween due to the relative rotation. When the external disturbance causes either one of the shaft 201 or sleeve 202 to deflect from the rotation axis, the force due to the dynamic pressure is offset to the rotation axis so that the horizontal component of the force revolves the rotational member around the rotation axis without returning the rotational member to its original position. The convergence of the revolution returns the rotational member to the original position so that the rotational member rotates in a stable manner. On the contrary, if the revolution is kept, the rotational member whirls around the central axis of the stationary member resulting in the unstable rotation. This phenomenon is referred to as the half-whirl. The present inventors have discovered that the revolution is likely to be kept with the bearing assembly having the continuous pressure distribution in comparison with one having a discontinuous pressure distribution.
FIG. 44 schematically illustrates the half-whirl phenomenon, showing a cross section along the rotation axis of the stationary shaft 201 and the rotating sleeve 202 of the hydrodynamic bearing assembly. In the normal operation, the sleeve 202 rotates around a rotation center concentric with the stationary axis I of the shaft 201 as indicated by the cross (+) in the direction of the arrow 215. When the external disturbance causes the sleeve 202 to deflect relatively to the shaft 201, the rotation center of the sleeve 202 of the rotational member is shifted from the stationary axis I to the position C as indicated by the alphabet (X). The force generated by the dynamic pressure having the deflecting direction as well as the continuous pressure distribution rotates the sleeve 202 on its own axis, and also revolves the rotation axis C of the sleeve 202 around the stationary axis I along the arrow 216 in a whirling manner. For example, the dashed line illustrates the sleeve 202a after the rotation axis C of the sleeve 202 indicated by the solid line revolves 180 degrees around the stationary axis I. In this instance, the rotation axis C of the sleeve 202 is shifted along the arrow 216 to the rotation axis C′. The half-whirl whirls the rotational member (such as sleeve 202 in FIG. 44) relatively to the stationary member (such as shaft 201 in FIG. 44) so that the bearing assembly loses the stability in rotation, thereby to cause undesired oscillation and/or malfunction of the bearing assembly used for the HDD or bar code reader.
(Second Problem)
The second problem to be solved, i.e., the rigidity/stiffness of the hydrodynamic bearing assembly will be discussed with reference of FIG. 45. This drawing is the enlarged view of the hydrodynamic bearing assembly of FIG. 42, in which similar reference numerals denote the similar components. In the drawing, the parallel lines schematically illustrate the dynamic pressure distribution generated during the rotation of the hydrodynamic bearing assembly. The dynamic pressure distribution M is generated in the radial bearing defined between the shaft 201 and the sleeve 202m, thereby to keep them away from each other. On the other hand, the dynamic pressure distribution N is also generated in the thrusts bearing defined between the thrust plate 203 and thrust opposing surface 204 so that no contact is kept therebetween, allowing the sleeve 202 of the rotational member to rotate without any contact.
The hydrodynamic bearing assembly of FIG. 45 receives external forces including a force indicated by the arrow 217 perpendicular to the bearing axis (translation force), a force indicated by the arrow 218 along the bearing axis (elevation force), a force indicated by the arrow 219 around an axis perpendicular to the bearing axis (oscillation force), and the combination thereof. The hydrodynamic bearing assembly is required to have a rigidity against such forces enough to keep the rotational member away from the stationary member and to ensure the stable rotation.
FIG. 46 provides an example where the sleeve 202 is inclined counterclockwise relative to the shaft 201 and the thrust plate 203 because of the disturbance (external forces) to the hydrodynamic bearing assembly during the stable rotation as shown in FIG. 45. In this instance, the shaft 201 moves closer to the sleeve at the right-upper portion indicated by T and at the left-lower portion indicated by U, also the thrust plate 203 moves closer to the thrust opposing surface 204 at the leftmost portion indicated by V. In general, the wedge effect due to the convolution of the fluid between relatively moving bearing members is increased as the gap therebetwen is decreased. Thus, the dynamic pressure distribution is shifted from as illustrated in FIG. 45 to that as shown in FIG. 46. The dynamic pressure is increased between the rotational member and the stationary member at the portions T and U so that the repulsion force is generated to prevent both members from moving closer to each other. The contact between the shaft 201 and the sleeve 202 is avoided unless the disturbance force overcomes the repulsion force.
Meanwhile, the fluid is guided from the circumference of thrust plate 203 towards the axis (the pump-in bearing assembly) so that the dynamic pressure between the thrust plate 203 and the thrust opposing surface 204 is increased towards the bearing axis, as shown by the portion N in FIG. 45. Thus, the peak of the dynamic pressure can disadvantageously be expected at the radially outer portion V shown in FIG. 46, even if the rotational member moves closer to the stationary member. Therefore, when the disturbance force causes oscillation force, the thrust plate 203 is likely to physically contact with the thrust opposing surface 204. Once the thrust plate 203 contacts with the thrust plate 203, the friction force therebetween results the unstable rotation of the rotation members. Further, the rebound followed by the contact causes the undesired impact, which could bring the malfunction of the magnetic head used for the HDD, or result an extensive damage to the spindle motor. Therefore, it is particularly important that the hydrodynamic bearing assembly has sufficient rigidity against the oscillation force, which is referred to as “tilt rigidity”. Also, the rigidity against the translation force and the rigidity against the elevation force, which are referred to as the “translation rigidity” and the “elevation rigidity”, respectively. Both of the translation rigidity and elevation rigidity can be improved by increasing the radial and thrust dynamic pressure.
FIG. 47 illustrates an exemplary hydrodynamic bearing assembly including the shaft 201 and the thrust plate 203 that is not perpendicularly attached thereto. The shaft 201 is inclined relative to the thrust plate 203, and the sleeve 202 is provided around the shaft 201. During the rotation of the spindle motor, in general, the radial bearing and the thrust bearing of the hydrodynamic bearing assembly have gaps of approximately 3 to 5 microns and approximately 2 to 10 microns, respectively, between the rotational member and the stationary member for the rotation without contact. In order to stabilize the rotation of the bearing assembly, the above-mentioned gaps are kept constant in a precise manner. In the radial bearing, since the inner surface of the sleeve 202 opposes to the outer surface of the shaft 201, the gap therebetween can readily be kept constant. On the other hand, the gap between the thrust opposing surface 204 of the sleeve 202 and the thrust plate 203 in the thrust bearing is more difficult to be kept constant than that in the radial bearing, because the gap in the thrust bearing is more susceptible to the arrangement of the sleeve 202 relative to the shaft 201. Thus, the precise control of the gap in the thrust bearing depends upon directly how the shaft 201 is arranged perpendicularly on the thrust plate 203 in a precise manner. Therefore, as illustrated in FIG. 47, in case where the shaft 201 is inclined to the thrust plate 203, the sleeve 202 rotate about the bearing axis inclined to the thrust plate 203, even if the normal dynamic pressure is generated in the radial bearing. This inclined rotation against the thrust plate 203 may raise a possibility that the sleeve 202 contacts with the thrust plate 203 in the portion V of FIG. 47 due to a slight oscillation force during the rotation.
(Third Problem)
The third problem to be solved is an improvement of the actuation feature of the hydrodynamic bearing assembly. When the hydrodynamic bearing assembly start to rotate, since no rotation generates no dynamic pressure, the sleeve 202 is in contact with the thrust plate 203, and in some cases, the shaft 201 also is in contact with the sleeve 202. Then, when the spindle motor is being actuated, the rotation at a relatively low rate keeps those members in contact with each other. The rotation rate exceeding to a predetermined rate generates the dynamic pressure enough to ensure the rotation without any contacts. This predetermined rotation rate is referred to as a “floating rotation rate”, hereinafter. Since the sleeve 202 rotates in contact with the thrust plate 203 before the floating rotation, there are problems of friction and overheat between the rotating and stationary member. Further, a greater driving torque is required to rotate the sleeve 202 in contact with the thrust plate 203. Thus, the higher floating rotation rate requires more time and energy consumption to achieve the rotation without any contacts. Therefore, the hydrodynamic bearing assembly has been demanded such that the floating rotation rate is minimized to rotate the rotational member keeping away from the stationary member within the shortest time in order to realize good endurance and less energy consumption for activation of the bearing assembly.
(Fourth Problem)
The fourth problem to be solved by the present invention is to realize the hydrodynamic bearing assembly to be more compact and lightweight. This need comes from the fact that devices such as memory device incorporating the hydrodynamic bearing assembly are demanded to be more compact and lightweight. Also, the more compact and lightweight bearing assembly advantageously causes the rotation with contact between the rotating and stationary member to wear less at the activation of the bearing assembly.
With respect to each of the problems to be solved as mentioned above, the prior art approaches to address the problems and the deficiencies thereof will independently be described hereinafter.
1. Half-Whirl
To address the problem of the half-whirl, the prior arts has proposed a plurality of notches provided parallel to the bearing axis on either one of the outer surface of the shaft 201 and the inner surface of the sleeve 202, which are opposing and rotates relative to the shaft 202. FIG. 48 is a vertical cross section, and FIG. 49 is a transverse cross section of the bearing assembly. As shown, the sleeve 202 is arranged around the shaft 201 for rotation of the sleeve 202 about the shaft 201. The shaft 201 is secured perpendicular onto the thrust plate 203, which opposes to the bottom surface of the sleeve 202. The hydrodynamic bearing assembly comprises the shaft 201, the sleeve 202, and the thrust plate 203.
As shown in FIG. 48, three longitudinal grooves 221 are formed on the outer surface along the bearing axis. By providing grooves 221, the continuous pressure distribution generated between the rotating and stationary member of the bearing assembly is interrupted to avoid the half-whirl phenomenon.
In general, taking account of the dynamic balance during high rotation, the grooves 221 are provided on the surface of the stationary member, which may be either one of outer surface of the shaft 201 and the inner surface of the sleeve 202. However, similar advantages can be enjoyed by providing the grooves on the surface of the rotational member. The radial dynamic pressure is reduced locally, and if the grooves are formed on the stationary member, then the translation rigidity is reduced along the direction of the arranged grooves. This approach may avoid the half-whirl but remains the disadvantage increasing the tendency to cause the rotational member in contact with the stationary member along the direction of the arranged grooves.
Another prior art has suggested providing either one of the outer surface of the shaft 201 and the inner surface of the sleeve 202 with a cross section of a configuration such as the triangle (the round-apex triangle) shape instead of the circle. This changes the gap between the shaft 201 and the sleeve 202 so that the aforementioned continuous pressure distribution is interrupted. For example, Japanese Patent Laid-Open Publication 02-89807 discloses the non-circular bearing assembly. Yet, the non-circular bearing assembly also causes the dynamic pressure in the broader gaps to reduce the translation rigidity.
Further, another prior art of Japanese Patent Laid-Open Publication 02-150504 discloses a plurality of longitudinal bands in the direction of the bearing axis including circumferential micro ground streaks formed on the inner surface of the sleeve 202. The ground streaks causes the turbulent flow of the fluid in the radial bearing so that the dynamic pressure distribution leading the half-whirl is prevented. FIGS. 50 and 51 are discussed in Japanese Patent Laid-Open Publication 02-150504. FIG. 50 illustrates the shaft 201 surrounded by the inner surface of the sleeve 202. FIG. 51 is an enlarged view of a portion W on the inner surface of the sleeve 202, in which the streak band 223 comprises circumferential micro ground streaks 222. A plurality of streak bands is formed in the direction of the bearing axis with a predetermined distance to each other. This overcomes the half-whirl while maintaining the translation rigidity, which might be reduced by the grooves formed on the inner surface of the sleeve 202. However, since each streak band 223 has to be scraped one by one, the task for scraping them is burdensome. In case where the inner diameter of the sleeve 202 has a small size, for example, in the order of several millimeters, scraping the streak bands 223 would be difficult.
2. Improvement of Rigidity of Bearing Assembly
Next, some conventional means having a main purpose for improving the rigidity of the bearing assembly will be described hereinafter. Japanese Patent Laid-Open Publication 08-338960 discloses an improvement of the radial rigidity by providing a plurality of shallow longitudinal grooves along the axis and an annular groove across the longitudinal grooves on the outer surface of the shaft of the hydrodynamic bearing assembly used for an optical scan device. This structure supports the sleeve on multi points against the radial motion so that the radial rigidity of the bearing assembly is improved. FIG. 52 is a vertical cross section of the optical scan device incorporating the hydrodynamic bearing assembly. In the hydrodynamic bearing assembly, the shaft 231 is mounted on the housing 230, and the shaft 231 is rotatably arranged around the shaft 231. A flange 233 made of metal such as aluminum and brass is secured on the outer surface. A polygonal mirror 234 for deflecting a laser beam is arranged on the top surface of the flange 233 by means of the spring 235. Also, a driving magnet 236 is bonded on the bottom surface at the perimeter of the flange 233 by the adhesive. A stator 238 is provided on the substrate 237 secured on the housing 230 so as to oppose to the driving magnet 236.
FIG. 53 is the enlarged view of the shaft 231 alone illustrating its detailed aspect and a schematic pattern of the dynamic pressure distribution (with its peaks). As shown, two of the parallel shallow grooves 241, 242 are formed on the surface for generating the dynamic pressure. Also, the annular groove 243 is formed on the surface of the shaft 231. The annular groove 243 divides the peak of the dynamic pressure distribution into two peaks Q1, Q2. This prevents the rotating sleeve 232 in contact with the shaft 231, thereby to avoid the damages to each other. It is understood that since the dynamic pressure distribution has two peaks, the bearing assembly with the annular groove 243 has greater anti-moment rigidity against the external moment than that without the annular groove. However, the dynamic pressure at the portions adjacent to the longitudinal grooves 241, 242 formed on the outer surface of the shaft 231 is reduced to decrease the translation rigidity at those portions. Also, other factors such as dimensions and weights of the annular grooves 243 further reduces the translation rigidity.
Another conventional technique has proposed biasing the rotational member in the radial bearing along a particular direction. This defines the minimum gap between the shaft and the sleeve at a predetermined point where the high dynamic pressure is generated. Thus, it is understood that the bearing rigidity is improved because of the high dynamic pressure generated on the point where the minimum gap is defined, and that the half-whirl can advantageously be avoided. This is true as well for a complex radial-thrust bearing assembly, in which the radial bearing assembly and the thrust bearing assembly are continuously formed.
In particular, according to Japanese Patent Laid-Open Publication 11-18357 and Japanese Patent Laid-Open Publication 11-55918, the rotor magnet is positioned eccentrically to the coil so that the shaft is biased against the sleeve in the predetermined direction for the stable rotation. FIG. 54 shows one example disclosed in Japanese Patent Laid-Open Publication 11-55918. The rotor 252 is arranged around the stator 251, the rotor magnet 253 attached to the inner surface of the rotor 252 is opposed to the stator 251 for driving the torque. As shown in FIG. 54, one stator 251a has an arm shorter than those of the remaining stators 251. This causes the gap h1 between the stator 251a and the rotor magnet 253 greater than the gap h2 between the stator 251 and the rotor magnet 253. Thus, the attraction force (or the repulsion force) is reduced in the gap h1 so that the rotor is biased against the stator in the predetermined direction. The stator 251 is secured concentrically to the shaft, the rotor 252 is biased against the shaft in the predetermined direction. However, in order to bias the shaft against the sleeve with use of the method, the stator 251 is required to be positioned concentrically to the baring axis. In practical, the alignment of the stator 251 is often impossible because of the design constraint.
Japanese Utility Model Laid-Open Publication 55-36456 discloses the stationary permanent magnet attached to the housing so as to oppose to the rotor magnet for tilting the rotor towards the predetermined direction for rotation. However, this approach causes the lifetime of the bearing assembly shorter because the distal edge of the shaft 255 contacts with the sleeve 254.
Some other prior arts use a pump-out type hydrodynamic bearing assembly to improve the bearing rigidity. The hydrodynamic bearing assembly, in which the fluid is conducted from the center towards the circumference, is referred to as the pump-out type hydrodynamic bearing assembly. Briefly speaking, the pump-out type hydrodynamic bearing assembly has spiral grooves 205 having the angle with the circle, or the rotating direction reversed to one indicated by the arrow 206 of FIG. 43. Since the pump-in type hydrodynamic bearing assembly has the dynamic pressure distribution with the peak adjacent to the axis, it is relatively susceptible to the external disturbance. Meanwhile, the pump-out type hydrodynamic bearing assembly has the peak of the dynamic pressure distribution at the outermost edge of the thrust plate 203, thereby to improve the rigidity against the disturbing motion.
FIGS. 55 and 56 illustrate one embodiment to implement the pump-out type hydrodynamic bearing assembly according to Japanese Patent Laid-Open Publication 9-229053. In FIG. 55, the sleeve 272 is arranged around the shaft 271. Also, the shaft 271 has the thrust plate 273, which is integrally formed and is flush with the surface perpendicular to the bearing axis. The shaft 271 and the thrust plate 273 together rotate within the chamber defined by the sleeve 272. The shaft 271 has herringbone grooves 274 on the outer surface.
FIG. 56 is a top view of the shaft 271 and the thrust plate 273 of FIG. 55. The thrust plate 273 includes a plurality of spiral grooves 275 for generating the dynamic pressure on both surfaces of the thrust plate 273 (including opposite surface of the drawing). The arrow 276 shows the rotational direction of the shaft 271 of the bearing assembly. As illustrated, the spiral grooves 275 on both surfaces of the thrust plate 273 are formed with the tilt so that they conduct the fluid within the bearing assembly radially from the center to the circumference. Also, the thrust plate 273 has a plurality of through-holes extending along the bearing axis adjacent to the shaft 271.
In FIG. 55, during the rotation of the bearing assembly, the herringbone grooves 274 guide the fluid away from the thrust plate 273. Contrary, the spiral grooves 275 of the pump-out type hydrodynamic bearing assembly conduct the fluid to the thrust bearing and the circumference of the thrust plate 273. To this end, the dynamic pressure distribution has a peak adjacent to the circumference of the thrust plate 273. The long and short dotted line 278 in FIG. 55 schematically shows the dynamic pressure distribution. The pump-out type hydrodynamic bearing assembly generates the peak dynamic pressure at the circumference so as to realize the robust rigidity against the disturbance motion when the fluid is supplied to the thrust member. The through-holes keeps the dynamic pressure above and under the thrust plate 273 even to stabilize the rotation of the bearing assembly.
As illustrated in FIG. 55, in the structure of the complex radial-thrust bearing assembly wherein the radial bearing and the thrust bearing are continuously formed, the thrust bearing requires the fluid supplied from the radial bearing to the thrust bearing, to be enough for generating the dynamic pressure. The fluid is supplied from the upper end of the shaft through the radial gap between the shaft 271 and the sleeve 272 during the rotation of the shaft 271. However, if the shaft 271 has a circular cross section as shown in FIG. 55, a sufficient amount of the fluid is hardly delivered to the trust plate 273. Also, the herringbone grooves 274 in the radial bearing generates a peak of the radial dynamic pressure distribution at the middle portion of the radial bearing (at the middle portion of the herringbones with the V-shaped indication in the drawing). The fluid for generating this peak at the middle portion of the herringbones comes towards the radial bearing so that the fluid supplied to the radial bearing is likely to be short even if the pump-out type spiral grooves 275 are provided on the thrust plate 273. When the fluid supplied to the thrust bearing is short, the dynamic pressure in the thrusting direction cannot be generated so that the supporting force is also short. To this end, this causes the rotational member and the stationary member to be in contact with each other.
The need has been existed for a simple fastening mechanism for perpendicularly fastening the shaft with the thrust plate in a precise manner, in order to improve the bearing rigidity. Some conventional fastening methods and problems thereof will be described hereinafter. The fastening mechanism may bond the shaft directly with the thrust plate. This mechanism has a difficulty to keep the accuracy of the perpendicularity due to an uneven thickness of the adhesive. The tolerance limits of the perpendicularity is 0.3 microns measured as the tilt relative to the thrust plate diameter of 20 millimeters. When the shaft has the diameter of 4 millimeters or more, it is hardly possible to meet the tolerance limits even if the adhesive is cured while held correctly.
As shown in FIG. 57(a), the fastening mechanism comprises a hollow cylindrical shaft 281, a thrust plate 283, and a volt extending therethrough for fastening the shaft and thrust plate. However, an uneven pressure biased by the bolt 286 or the washer 287 causes the elastic deformation of the shaft 281, thereby to result the malfunction of the bearing assembly in this mechanism. Even a rubber pad attached to the end of the hollow space of the shaft 281 provides the same result. The present inventors have found that the perpendicularity was 1.2 micron due to the uneven pressure biased by the bolt 286 and it was 1.0 micron with use of the rubber pad. It is hardly possible to reduce the deviation of the radial component of the fastening force with use of this mechanism.
Also, another fastening mechanism comprises the cylindrical hollow shaft 281, the thrust plate 283′, and a core member 288 secured on the thrust plate 283′ and connected with the shaft 281 by the shrink fitting. However, when cooled down to the room temperature, the outer surface of the shaft 281 radially expands with the elastic deformation so that the radial gap between the shaft and the sleeve has an unwanted influences. The present inventors have found that the outer diameter of the shaft 281 radially expanded by 3 microns due to the shrink fitting and the bearing assembly could not practically be used.
Another prior arts fastening mechanism fastens the shaft and thrust plate by screwing a bolt onto the shaft 281 and the thrust plate 283. This fastening mechanism is simple and keeps the perpendicularity in a precise manner. If the shaft 281 is made of stainless steel, then this fastening mechanism can be used. However, if the shaft 281 is made of ceramics material, in which it is difficult to make a thread, the mechanism can hardly be utilized in practical.
(Third Problem)
To address the third problem, i.e., the improvement of the activation feature, many approaches have been proposed, for example by developing an effective spiral grooves in generating the thrust dynamic pressure. One of the solutions is reducing the floating rotation rate. Other solutions include increasing the acceleration at the beginning of the activation to minimize the time period in rotating with contact, and reducing the mass of the rotational member. Unless the rotation with contact is avoided, at the beginning of the activation, a significant activation torque is required and the bearing assembly wears quickly. Thus, the activation feature has to be further improved in future.
(Fourth Problem)
With respect to the fourth problem, i.e., an implementation of the compact and lightweight bearing assembly, the hydrodynamic bearing assembly has achieved the improvement in comparison with the ball bearing assembly. However, in any event, the market still needs the bearing assembly to be more compact and lightweight, thus, a further improvement is required.
Problems to be Solved
The conventional approaches for overcoming the half-whirl of the bearing assembly and for improving the bearing rigidity have each deficiencies as described above. Therefore, a purpose of the present invention is to provide the hydrodynamic bearing assembly eliminating the half-whirl and improving the rigidity against the disturbance.
Another purpose of the present invention is to provide the hydrodynamic bearing assembly improving the activation feature and satisfying the compact and lightweight requirement.